The present invention relates to the automotive timing chain art. It finds particular application in conjunction with a roller chain sprocket for use in automotive camshaft drive applications and will be described with particular reference thereto. However, the present invention may also find application in conjunction with other types of chain drive systems and applications where reducing the noise levels associated with chain drives is desired.
Roller chain sprockets for use in camshaft drives of automotive engines are typically manufactured according to one or more international standards such as DIN, JIS, ISO, etc. The ISO-606:1994(E) (International Organization for Standardization) standard specifies requirements for short-pitch precision roller chains and associated chain wheels or sprockets.
FIG. 1 illustrates a symmetrical tooth space form for an ISO-606 compliant sprocket. The tooth space has a continuous fillet or root radius Ri extending from one tooth flank (i.e., side) to the adjacent tooth flank as defined by the roller seating angle α. The flank radius Rf is tangent to the roller seating radius Ri at the tangency point TP. A chain with a link pitch P has rollers of diameter D1 in contact with the tooth spaces. The ISO sprocket has a chordal pitch also of length P, a root diameter D2, and Z number of teeth. The pitch circle diameter PD, tip or outside diameter OD, and tooth angle A (equal to 360°/Z) further define the ISO-606 compliant sprocket. The maximum and minimum roller seating angle α is defined as:αmax=140°−(90°/Z) and αmin=120°−(90°/Z)
With reference to FIG. 2, an exemplary ISO-606 compliant roller chain drive system 10 rotates in a clockwise direction as shown by arrow 11. The chain drive system 10 includes a drive sprocket 12, a driven sprocket 14 and a roller chain 16 having a number of rollers 18. The sprockets 12, 14, and chain 16 each generally comply with the ISO-606 standard.
The roller chain 16 engages and wraps about sprockets 12 and 14 and has two spans extending between the sprockets, slack strand 20 and taut strand 22. The roller chain 16 is under tension as shown by arrows 24. A central portion of the taut strand 22 may be guided between the driven sprocket 14 and the drive sprocket 12 with a conventional chain guide (not shown). A first roller 28 is shown at the onset of meshing at a 12 o'clock position on the drive sprocket 12. A second roller 30 is adjacent to the first roller 28 and is the next roller to mesh with the drive sprocket 12.
Chain drive systems have several components of undesirable noise. A major source of roller chain drive noise is the sound generated as a roller leaves the span and collides with the sprocket during meshing. The resultant impact noise is repeated with a frequency generally equal to that of the frequency of the chain meshing with the sprocket. The loudness of the impact noise is a function of the impact energy (EA) occurring during the meshing process. The impact energy (EA) is related to engine speed, chain mass, and the impact velocity between the chain and the sprocket at the onset of meshing. The impact velocity is affected by the chain-sprocket engagement geometry, of which an engaging flank
                    E        A            =                        wP          2000                ⁢                  V          A          2                      ;                      V        A            =                                    π            ⁢                                                  ⁢            nP                    30000                ⁢                  sin          ⁡                      (                                          360                Z                            +              γ                        )                                ;              γ      =                        180          -          A          -          α                2              ;    and  pressure angle γ (FIG. 3) is a factor, where:                EA=Impact Energy [N·m]        VA=Roller Impact Velocity [m/s]        γ=Engaging Flank Pressure Angle        n=Engine Speed [RPM]        w=Chain Mass [Kg/m]        Z=Number of Sprocket Teeth        A=Tooth Angle (360°/Z)        α=Roller Seating Angle        P=Chain Pitch (Chordal Pitch)        
The impact energy (EA) equation presumes the chain drive kinematics will conform generally to a quasi-static analytical model and that the roller-sprocket driving contact will occur at a tangent point TP (FIG. 3) of the flank and root radii as the sprocket collects a roller from the span.
As shown in FIG. 3, the pressure angle γ is defined as the angle between a line A extending from the center of the engaging roller 28, when it is contacting the engaging tooth flank at the tangency point TP, through the center of the flank radius Rf, and a line B connecting the centers of the fully seated roller 28, when it is seated on root diameter D2, and the center of the next meshing roller 30, as if it were also seated on root diameter D2 in its engaging tooth space. The roller seating angles α and pressure angles γ listed in FIG. 4 are calculated from the equations defined above. It should be appreciated that γ is a minimum when α is a maximum. Thus, the exemplary 23-tooth, ISO-606 compliant, drive sprocket 12 shown in FIGS. 2 and 3 will have a pressure angle γ in the range of 14.13° to 24.13° as listed in the table of FIG. 4.
FIG. 3 also shows the engagement path (phantom rollers) and the driving contact position of roller 28 (solid) as the drive sprocket 12 rotates in the direction of arrow 11. FIG. 3 depicts the theoretical case with chain roller 28 seated on root diameter D2 of a maximum material sprocket with both chain pitch and sprocket chordal pitch equal to theoretical pitch P. For this theoretical case, the noise occurring at the onset of roller engagement has a radial component FR as a result of roller 28 colliding with the root surface Ri and a tangential component FT generated as the same roller 28 collides with the engaging tooth flank at point TP as the roller moves into driving contact. It is believed that the radial impact occurs first, with the tangential impact following nearly simultaneously. Roller impact velocity VA is shown to act through, and is substantially normal to, engaging flank tangent point TP with roller 28 in driving contact at point TP.
The impact energy (EA) equation accounts only for a tangential roller impact during meshing. The actual roller engagement, presumed to have a tangential and radial impact (occurring in any order), would therefore seem to be at variance with the impact energy (EA) equation. The application of this quasi-static model, which is beneficially used as a directional tool, permits an analysis of those features that may be modified to reduce the impact energy occurring during the tangential roller-sprocket collision at the onset of meshing. The radial collision during meshing, and its effect on noise levels, can be evaluated apart from the impact energy (EA) equation.
Under actual conditions as a result of feature dimensional tolerances, there will normally be a pitch mismatch between the chain and sprocket, with increased mismatch as the components wear in use. This pitch mismatch serves to move the point of meshing impact, with the radial collision still occurring at the root surface Ri but not necessarily at D2. The tangential collision will normally be in the proximity of point TP, but this contact could take place high up on the engaging side of root radius Ri or even radially outward from point TP on the engaging flank radius Rf as a function of the actual chain-sprocket pitch mismatch.
Reducing the engaging flank pressure angle γ reduces the meshing noise levels associated with roller chain drives, as predicted by the impact energy (EA) equation set forth above. It is feasible but not recommended to reduce the pressure angle γ while maintaining a symmetrical tooth profile, which could be accomplished by simply increasing the roller seating angle α, effectively decreasing the pressure angle for both flanks. This profile as described requires that a worn chain would, as the roller travels around a sprocket wrap (discussed below), interface with a much steeper incline and the rollers would necessarily ride higher up on the coast flank prior to leaving the wrap.
Another source of chain drive noise is the broadband mechanical noise generated in part by shaft torsional vibrations and slight dimensional inaccuracies between the chain and the sprockets. Contributing to a greater extent to the broadband mechanical noise level is the intermittent or vibrating contact that occurs between the unloaded rollers and the sprocket teeth as the rollers travel around the sprocket wrap. In particular, ordinary chain drive system wear comprises sprocket tooth face wear and chain wear. The chain wear is caused by bearing wear in the chain joints and can be characterized as pitch elongation. It is believed that a worn chain meshing with an ISO standard sprocket will have only one roller in driving contact and loaded at a maximum loading condition.
With reference again to FIG. 2, driving contact at maximum loading occurs as a roller enters a drive sprocket wrap 32 at engagement. Engaging roller 28 is shown in driving contact and loaded at a maximum loading condition. The loading on roller 28 is primarily meshing impact loading and the chain tension loading. The next several rollers in the wrap 32 forward of roller 28 share in the chain tension loading, but at a progressively decreasing rate. The loading of roller 28 (and to a lesser extent for the next several rollers in the wrap) serves to maintain the roller in solid or hard contact with the sprocket root surface 34.
A roller 36 is the last roller in the drive sprocket wrap 32 prior to entering the slack strand 20. Roller 36 is also in hard contact with drive sprocket 12, but at some point higher up (e.g., radially outwardly) on the root surface 34. With the exception of rollers 28 and 36, and the several rollers forward of roller 28 that share the chain tension loading, the remaining rollers in the drive sprocket wrap 32 are not in hard contact with the sprocket root surface 34, and are therefore free to vibrate against the sprocket root surfaces as they travel around the wrap, thereby contributing to the generation of unwanted broadband mechanical noise.
A roller 38 is the last roller in a sprocket wrap 40 of the driven sprocket 14 before entering the taut strand 22. The roller 38 is in driving contact with the sprocket 14. As with the roller 36 in the drive sprocket wrap 32, a roller 42 in the sprocket wrap 40 is in hard contact with a root radius 44 of driven sprocket 14, but generally not at the root diameter.
It is known that providing pitch line clearance (PLC) between sprocket teeth promotes hard contact between the chain rollers and sprocket in the sprocket wrap, even as the roller chain wears. The amount of pitch line clearance added to the tooth space defines a length of a short arc that is centered in the tooth space and forms a segment of the root diameter D2. The root fillet radius Ri is tangent to the flank radius RF and the root diameter arc segment. The tooth profile is still symmetrical, but Ri is no longer a continuous fillet radius from one flank radius to the adjacent flank radius. This has the effect of reducing the broadband mechanical noise component of a chain drive system. However, adding pitch line clearance between sprocket teeth does not reduce chain drive noise caused by the roller-sprocket collision at impact.
Another attempt to reduce the noise levels associated with roller chain meshing is described in U.S. Pat. No. 5,397,278 which discloses the undercutting or relieving of the root surfaces so as to eliminate the radial roller-root surface contact at the onset of meshing. However, the invention disclosed in the '278 patent does not modulate the meshing impact frequency. That is, all of the tooth profiles are substantially identical. Therefore, the flank impacts occur at the meshing frequency. An additional disadvantage of the sprocket disclosed in the '278 patent is that the rollers contact both engaging and disengaging flanks at full mesh. Thus, a roller can become wedged within the tooth space when no clearance is provided between the roller and a disengaging flank with the roller seated in full mesh.
A further attempt to reduce the noise levels associated with roller chain meshing is to incorporate one or more elastomeric cushion rings that serve to buffer or soften the engaging impact of a roller as it leaves the span and collides with a sprocket during the meshing process. With reference to FIG. 6, a drive sprocket 112, associated with a conventional roller chain drive system 110, incorporates symmetrical, ISO-606 compliant tooth space profiles. The drive sprocket 112 is substantially identical to the drive sprocket 12 (FIG. 2) except that the drive sprocket 112 includes two circular cushion rings 144, one secured on each hub of the sprocket 112. Each cushion ring 144 has a continuous or otherwise uniform outer surface that is defined by a radius R. As is known in the art, the cushion rings 144 serve to buffer or soften the engaging impact of a roller as it leaves the span and collides with a sprocket during the meshing process.
More particularly, as shown in FIGS. 6a and 6b, a roller 128, a bushing 129, and an associated pin 131, are carried by two sets of overlapping link plates 146L, 146R and 148L, 148R. As the drive sprocket 112 rotates in the direction of arrow 11, the link plates 146 impact, and then compress or otherwise deform, the outer surfaces of both cushion rings 144 prior to roller 128 colliding with the associated sprocket tooth, followed consecutively by the link plates 148. As a result, the impact velocity of the roller 128 is reduced prior to meshing with the sprocket, thereby reducing meshing impact noise. The maximum amount of cushion ring compression 149 occurs close to, or at, the midpoint P/2 between adjacent rollers having a chain pitch P. FIG. 6b shows the rubber compression to alternate in consecutive pitches between link plates 146 and link plates 148. A typical problem with cushion rings is one of durability. That is, wear and compression set of the elastomeric material of the cushion rings can be expected in the area of maximum ring compression after repeated link impacts. In addition, the cushion ring may eventually fatigue as a result of the repeated compression cycles during the meshing process.
Accordingly, it is considered desirable to develop a new and improved roller chain drive system and sprocket which meets the above-stated needs and overcomes the foregoing disadvantages and others while providing better and more advantageous results.